Hybrid power transmission system having first and second clutch mechanisms

BACKGROUND OF THE INVENTION

The present invention relates to a vehicular rotational apparatus that has a rotary body and an electric motor. The rotary body is operably connected to a rotary shaft, which drives a mechanism, and transmits power to the rotary shaft from an external drive source. The electric motor selectively drives the rotary shaft.

A typical compressor drives a compressing mechanism for compressing refrigerant by selectively using power from an external drive source and power from an electric motor, which is located on the compressor. Japanese Laid-Open Patent Publication No. 11-30182 discloses such compressor.

The compressor of the above publication has a pulley for receiving power from the external drive source and a rotary shaft for driving the compression mechanism. A pulley one-way clutch is located in a power transmission path between the pulley and the rotary shaft. A motor one-way clutch is located in a power transmission path between the electric motor for driving the compression mechanism and the rotary shaft.

Therefore, the compression mechanism is driven by power from the external drive force without rotating the rotor of the electric motor. As a result, power transmitted from the external drive source to the rotary shaft is prevented from being consumed unnecessarily for purposes other than driving compression mechanism.

In the above structure, the size of the compressor is reduced by using the one-way clutch instead of an electromagnetic clutch. However, the location of, for example, the electric motor to reduce the size of the compressor is not disclosed.

SUMMARY OF THE INVENTION

Accordingly, it is an objective of the present invention to provide a vehicular rotational apparatus that is minimized in the axial direction of a rotary shaft.

To achieve the above objective, the present invention provides a vehicular rotational apparatus that is driven by an external drive source. The rotational apparatus includes a housing, a rotary shaft, a first rotary body, an electric motor, a first clutch mechanism, and a second clutch mechanism. The rotary shaft is rotatably supported by the housing. The first rotary body is operably connected to the rotary shaft and includes a power transmission portion. When power is transmitted from the external drive source to the power transmission portion, the first rotary body is rotated. The electric motor rotates the rotary shaft and includes a second rotary body, which is operably connected to the rotary shaft. At least part of the electric motor overlaps the power transmission portion in the axial direction of the rotary shaft. The first clutch mechanism is located between the first rotary body and the rotary shaft and selectively permits and prevents power transmission between the first rotary body and the rotary shaft. The second clutch mechanism is located between the second rotary body and the rotary shaft and selectively permits and prevents power transmission between the second rotary body and the rotary shaft.

Other aspects and advantages of the invention will become apparent from the following description, taken in conjunction with the accompanying drawings, illustrating by way of example the principles of the invention.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention, together with objects and advantages thereof, may best be understood by reference to the following description of the presently preferred embodiments together with the accompanying drawings in which:

FIG. 1

is a schematic cross-sectional view illustrating a compressor according to a first embodiment of the present invention;

FIG. 2

is a schematic cross-sectional view illustrating a control valve according to the first embodiment;

FIGS.

3

(

a

) and

3

(

b

) is an enlarged partial cross-sectional view illustrating a clutch according to the first embodiment;

FIG.

4

(

a

) is a front view illustrating a power transmission mechanism according to a second embodiment;

FIG.

4

(

b

) is a cross-sectional view taken along line

4

b

—

4

b

in FIG.

4

(

a

);

FIG. 5

is an enlarged partial cross-sectional view illustrating rubber dumpers and a power transmission piece according to the second embodiment; and

FIG. 6

is an enlarged cross-sectional view illustrating a power transmission mechanism according to a third embodiment of the present invention.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

A compressor C according to a first embodiment of the present invention will now be described with reference to

FIGS. 1

to

3

(

b

). The left end of the compressor C in

FIG. 1

is defined as the front of the compressor, and the right end is defined as the rear of the compressor C.

The compressor C forms a part of a vehicular air conditioner. As shown in

FIG. 1

, the compressor C includes a cylinder block

11

, a front housing member

12

, a valve plate assembly

13

, and a rear housing member

14

. The front housing member

12

is secured to the front end of the cylinder block

11

. The rear housing member

14

is secured to the rear end of the cylinder block

11

with the valve plate assembly

13

in between. The cylinder block

11

, the front housing member

12

, the valve plate assembly

13

, and the rear housing member

14

form the housing of the compressor C.

The cylinder block

11

and the front housing member

12

define a control pressure zone, which is a crank chamber

15

in the first embodiment, in between.

A rotary shaft

16

is housed in the compressor housing and extends through the crank chamber

15

. The front portion of the rotary shaft

16

is supported by a radial bearing

12

A located in the front wall of the front housing member

12

. The rear portion of the rotary shaft

16

is supported by a radial bearing

11

A located in the cylinder block

11

.

The front end portion of the rotary shaft

16

extends through the front wall of the front housing member

12

. A power transmission mechanism PT is fixed to the front end of the rotary shaft

16

. The power transmission mechanism PT includes a first rotary body, which is a pulley

17

in the first embodiment. The front end of the rotary shaft

16

is coupled to an external drive source, which is a vehicular engine E in the first embodiment, by the power transmission mechanism PT and a belt

18

, which is engaged with the pulley

17

.

A sealing member

12

B is located between the front end of the rotary shaft

16

and the front wall of the front housing member

12

. The sealing member

12

B is located outward of the front housing member

12

from the radial bearing

12

A in the axial direction of the rotary shaft

16

. The sealing member

12

B separates the inside and outside of the compressor housing.

The power transmission mechanism PT and the compressor C form a vehicular rotational apparatus in the first embodiment.

A lug plate

19

is coupled to the rotary shaft

16

and is located in the crank chamber

15

. The lug plate

19

rotates integrally with the rotary shaft

16

. A drive plate, which is a swash plate

20

in the first embodiment, is housed in the crank chamber

15

. The swash plate

20

slides along and inclines with respect to the rotary shaft

16

. The swash plate

20

is coupled to the lug plate

19

by a hinge mechanism

21

. The lug plate

19

permits the swash plate

20

to rotate integrally with the rotary shaft

16

and to incline with respect to the rotary shaft

16

while sliding along the rotation axis of the rotary shaft

16

.

A snap ring

22

is fitted about the rotary shaft

16

. A spring

23

extends between the snap ring

22

and the swash plate

20

. The snap ring

22

and the spring

23

limit the minimum inclination angle of the swash plate

20

. At the minimum inclination angle of the swash plate

20

, the angle defined by the swash plate

20

and the axis of the rotary shaft

16

is closest to ninety degrees.

Cylinder bores

24

(only one is shown in

FIG. 1

) are formed in the cylinder block

11

. The cylinder bores

24

are located about the rotation axis of the rotary shaft

16

. A single-headed piston

25

is housed in each cylinder bore

24

to reciprocate inside the cylinder bore

24

. The front and rear openings of each cylinder bore

24

are closed by the associated piston

25

and the valve plate assembly

13

. A compression chamber is defined in each cylinder bore

24

. The volume of the compression chamber changes according to the reciprocation of the corresponding piston

25

. Each piston

25

is coupled to the peripheral portion of the swash plate

20

by a pair of shoes

26

. When the swash plate

20

is rotated by rotation of the rotary shaft

16

, the shoes

26

convert the rotation into reciprocation of each piston

25

.

The cylinder block

11

, the rotary shaft

16

, the lug plate

19

, the swash plate

20

, the hinge mechanism

21

, the pistons

25

, and the shoes

26

form a piston type variable displacement compression mechanism. The compression mechanism is driven by the rotary shaft

16

to compress and discharge refrigerant.

A suction pressure zone, which is a suction chamber

27

in the first embodiment, and a discharge pressure zone, which is a discharge chamber

28

in the first embodiment, are defined in the rear housing member

14

. The front ends of the suction chamber

27

and the discharge chamber

28

are closed by the valve plate assembly

13

. As each piston

25

moves from the top dead center position to the bottom dead center position, refrigerant gas is drawn into the corresponding cylinder bore

24

(compression chamber) through the corresponding suction port

29

while flexing the suction valve flap

30

to an open position. Low pressure refrigerant gas that is drawn into the cylinder bore

24

is compressed to a predetermined pressure as the piston

25

is moved from the bottom dead center position to the top dead center position. Then, the gas is discharged to the discharge chamber

28

through the corresponding discharge port

31

while flexing the discharge valve flap

32

to an open position.

The suction chamber

27

is connected to the discharge chamber

28

by an external refrigerant circuit

33

. The external refrigerant circuit

33

includes a condenser

34

, a decompression device, which is an expansion valve

35

in the first embodiment, and an evaporator

36

. The opening degree of the expansion valve

35

is feedback-controlled based on the temperature and pressure of refrigerant detected by a heat sensitive tube (not shown) at the outlet, or downstream, of the evaporator

36

. The expansion valve

35

supplies refrigerant, the amount of which corresponds to the thermal load (cooling load), to the evaporator

36

to regulate the flow rate in the external refrigerant circuit

33

.

A connecting pipe

37

for refrigerant gas is located at a downstream portion of the external refrigerant circuit

33

and connects the outlet of the evaporator

36

to the suction chamber

27

of the compressor C. Another connecting pipe

38

for refrigerant gas is located at an upstream portion of the external refrigerant circuit

33

and connects the discharge chamber

28

of the compressor C to the condenser

34

. The compressor C draws in refrigerant gas introduced into the suction chamber

27

from the downstream portion of the external refrigerant circuit

33

and compresses the refrigerant gas. Then, the compressor C discharges the compressed gas to the discharge chamber

28

, which is connected to the upstream portion of the external refrigerant circuit

33

.

The compressor C and the external refrigerant circuit

33

constitute a refrigeration circuit (or refrigerant circuit) of the vehicular air-conditioner.

The cylinder block

11

has a shaft bore

39

, which accommodates the rear end of the rotary shaft

16

. A shaft passage

40

is formed in the rotary shaft

16

to connect the front portion of the crank chamber

15

with the shaft bore

39

. A communication passage

41

is formed in the valve plate assembly

13

to connect the suction chamber

27

with the shaft bore

39

. The shaft bore

39

, the shaft passage

40

, and the communication passage

41

constitutes a bleed passage, which connects the crank chamber

15

with the suction chamber

27

.

A supply passage

42

is formed in the housing to connect the discharge chamber

28

with the crank chamber

15

. A control valve

43

is located in the supply passage

42

to regulate the opening degree of the supply passage

42

.

The opening of the control valve

43

is adjusted to control the flow rate of highly pressurized gas supplied to the crank chamber

15

through the supply passage

42

. The pressure in the crank chamber

15

(crank chamber pressure Pc) is determined by the ratio of the gas supplied to the crank chamber

15

through the supply passage

42

and the flow rate of refrigerant gas conducted out from the crank chamber

15

through the bleed passage. As the crank chamber pressure Pc varies, the difference between the crank chamber pressure Pc and the pressure in the compression chambers varies, which changes the inclination angle of the swash plate

20

. Accordingly, the amount of refrigerant discharged from the compression mechanism during one rotation of the rotary shaft

16

is varied.

The compressor C according to the first embodiment is formed such that the amount of refrigerant discharged during one rotation of the rotary shaft

16

is substantially zero when the inclination angle of the swash plate

20

is minimum.

The greater the flow rate of the refrigerant Q flowing in the refrigerant circuit is, the greater the pressure loss per unit length of the circuit or piping is. That is, the pressure loss (pressure difference) between first and second pressure monitoring points P

1

, P

2

has a positive correlation with the flow rate of the refrigerant Q in the refrigerant circuit. Detecting the pressure difference between the first and second pressure monitoring points P

1

, P

2

(PdH−PdL=pressure difference &Dgr;PX) permits the flow rate of refrigerant Q in the refrigerant circuit to be indirectly detected.

In the first embodiment, the first pressure monitoring point P

1

, which functions as a high pressure monitoring point, is located in the discharge chamber

28

, the pressure of which is equal to that of the most upstream section of the connecting pipe

38

. The second pressure monitoring point P

2

, which functions as a low pressure monitoring point, is located midway along the connecting pipe

38

at a position separated from the first pressure monitoring point P

1

by a predetermined distance. The pressure PdH at the first pressure monitoring point P

1

is applied to the control valve

43

through a first pressure introduction passage

44

(see FIG.

2

). The pressure PdL at the second pressure monitoring point P

2

is applied to the control valve

43

through a second pressure introduction passage

45

(see FIG.

2

).

A throttle

46

may be formed in the connecting pipe

38

between the first and second pressure monitoring points P

1

, P

2

to increase the pressure difference &Dgr;PX. The throttle

46

increases the pressure difference &Dgr;PX between the first and second pressure monitoring points P

1

, P

2

although the first and second pressure monitoring points P

1

, P

2

are not separated by a large amount. Providing the throttle

46

between the first and second pressure monitoring points P

1

, P

2

permits the second pressure monitoring point P

2

to be located close to the compressor C. This shortens the second pressure introduction passage

45

between the second pressure monitoring point P

2

and the control valve

43

. The pressure PdL at the second pressure monitoring point P

2

is set sufficiently higher than the crank chamber pressure Pc although the pressure PdL is decreased with respect to the pressure PdH by the throttle

46

.

As shown in

FIG. 2

, the control valve

43

has a valve housing

47

. The valve housing

47

defines a valve chamber

48

, a communication passage

49

, and a pressure sensing chamber

50

. A transmission rod

51

extends through the valve chamber

48

and the communication passage

49

. The transmission rod

51

moves in the axial direction, or in the vertical direction as viewed in FIG.

2

.

The communication passage

49

is disconnected from the pressure sensing chamber

50

by the upper portion of the transmission rod

51

, which is fitted in the communication passage

49

. The valve chamber

48

is connected to the discharge chamber

28

through an upstream section of the supply passage

42

. The communication passage

49

is connected to the crank chamber

15

by a downstream section of the supply passage

42

. The valve chamber

48

and the communication passage

49

form a part of the supply passage

42

.

A valve body

52

is formed at the middle portion of the transmission rod

51

and is located in the valve chamber

48

. A step defined between the valve chamber

48

and the communication passage

49

functions as a valve seat

53

. The communication passage

49

serves as a valve hole. The transmission rod

51

shown in

FIG. 2

is located at the lowermost position where the opening degree of the communication passage

49

is the greatest. When the transmission rod

51

is moved from the lowermost position to the uppermost position, at which the valve body

52

contacts the valve seat

53

, the communication passage

49

is disconnected from the valve chamber

48

. That is, the valve body

52

of the transmission rod

51

is a valve body that controls the opening degree of the supply passage

42

.

A pressure sensing member, which is a bellows

54

in the first embodiment, is located in the pressure sensing chamber

50

. The upper end of the bellows

54

is fixed to the valve housing

47

. The lower end of the bellows

54

receives the upper end of the transmission rod

51

. The bellows

54

divides the pressure sensing chamber

50

into a first pressure chamber

55

, which is the interior of the bellows

54

, and a second pressure chamber

56

, which is the exterior of the bellows

54

. The first pressure chamber

55

is connected to the first pressure monitoring point P

1

by the first pressure introduction passage

44

. The second pressure chamber

56

is connected to the second pressure monitoring point P

2

by the second pressure introduction passage

45

. Therefore, the first pressure chamber

55

is exposed to the pressure PdH monitored at the first pressure monitoring point P

1

, and the second pressure chamber

56

is exposed to the pressure PdL monitored at the second pressure monitoring point P

2

. The bellows

54

and the pressure sensing chamber

50

form a pressure sensing mechanism.

A target pressure difference changing means, which is an electromagnetic actuator

57

in the first embodiment, is located at the lower portion of the valve housing

47

. The electromagnetic actuator

57

includes a cup-shaped cylinder

58

, which is arranged coaxial to the valve housing

47

. A stationary iron core

59

is fitted in the upper opening of the cylinder

58

and is secured to the cylinder

58

. The stationary iron core

59

defines a plunger chamber

60

at the lowermost portion in the cylinder

58

.

A movable iron core

61

is located in the plunger chamber

60

. The movable iron core

61

slides along the plunger chamber

60

in the axial direction. An axially extending guide hole

62

is formed in the central portion of the stationary iron core

59

. The lower end of the transmission rod

51

is located in the guide hole

62

to move axially. The lower end of the transmission rod

51

abuts against the movable iron core

61

in the plunger chamber

60

.

A coil spring, which is a spring

63

in the first embodiment, is located between the inner bottom surface of the cylinder

58

and the movable iron core

61

in the plunger chamber

60

. The spring

63

urges the movable iron core

61

toward the transmission rod

51

. The transmission rod

51

is urged toward the movable iron core

61

by the elasticity of the bellows

54

. Therefore, the movable iron core

61

and the transmission rod

51

integrally move vertically. Hereinafter, urging force based on the elasticity of the bellows

54

is referred to as the spring force of the bellows. The spring force of the bellows

54

is greater than the force of the spring

63

.

A coil

64

is wound about the stationary iron core

59

and the movable iron core

61

on the outer circumference of the cylinder

58

. Power is supplied to the coil

64

from a battery via a drive circuit (not shown) based on commands from a controller, which is not shown.

The coil

64

generates an electromagnetic force (electromagnetic attracting force) between the movable iron core

61

and the stationary iron core

59

in accordance with the value of current supply to the coil

64

. Upward force is applied to the transmission rod

51

via the movable iron core

61

in accordance with the electromagnetic force. In the first embodiment, current supplied to the coil

64

is varied by controlling the applied voltage. The applied voltage is controlled by pulse-width modulation, or duty control.

According to the control valve

43

, the position of the transmission rod

51

(valve body

52

), or the opening degree, is determined in the following manner.

When no current is supplied to the coil

64

, or when the duty ratio is zero percent, the downward force generated by the spring force of the bellows

54

dominantly acts on the transmission rod

51

. Thus, the transmission rod

51

is placed at its lowermost position, and the communication passage

49

is fully opened. Therefore, the crank chamber pressure Pc is the maximum that is possible under the given conditions. The pressure difference between the crank chamber pressure Pc and the pressure in the compression chambers thus becomes large. As a result, the inclination angle of the swash plate

20

is minimized and the amount of refrigerant discharged during one rotation of the rotary shaft

16

is also minimized.

When a current of the minimum duty ratio or more within the variation range of the duty ratio is supplied to the coil

64

, the resultant of the upward force of the spring

63

and the upward electromagnetic force exceeds the downward force generated by the spring force of the bellows

54

so that the transmission rod

51

is moved upward. In this state, the resultant of the upward force of the spring

63

and the upward electromagnetic force acts against the resultant of the force based on the pressure difference &Dgr;PX and the downward force generated by the spring force of the bellows

54

. The position of the valve body

52

of the transmission rod

51

relative to the valve seat

53

is determined such that upward and downward forces are balanced.

For example, if the flow rate of refrigerant in the refrigerant circuit is decreased, the downward force based on the pressure difference &Dgr;PX, which acts on the transmission rod

51

, decreases. Therefore, the transmission rod

51

(valve body

52

) moves upward to decrease the opening degree of the communication passage

49

, which lowers the crank chamber pressure Pc. Accordingly, the inclination angle of the swash plate

20

is increased, and the displacement of the compressor C is increased. The increase in the displacement of the compressor C increases the flow rate of refrigerant in the refrigerant circuit, which increases the pressure difference &Dgr;PX.

In contrast, when the flow rate of refrigerant in the refrigerant circuit is increased, the downward force based on the pressure difference &Dgr;PX increases. Therefore, the transmission rod

51

(valve body

52

) moves downward to increase the opening degree of the communication passage

49

, which increases the crank chamber pressure Pc. Accordingly, the inclination angle of the swash plate

20

is decreased, and the displacement of the compressor C is decreased. The decrease in the displacement of the compressor C decreases the flow rate of refrigerant in the refrigerant circuit, which decreases the pressure difference &Dgr;PX.

When the duty ratio of the electric current supplied to the coil

64

is increased to increase the upward electromagnetic force, the pressure difference &Dgr;PX cannot balance the forces acting on the transmission rod

51

. Therefore, the transmission rod

51

(the valve body

52

) moves upward and decreases the opening degree of the communication passage

49

. As a result, the displacement of the compressor C is increased. Thus, the flow rate of refrigerant in the refrigerant circuit increases, which increases the pressure difference &Dgr;PX.

When the duty ratio of the electric current supplied to the coil

64

is decreased to decrease the upward electromagnetic force, the pressure difference &Dgr;PX cannot balance the forces acting on the transmission rod

51

. Therefore, the transmission rod

51

(the valve body

52

) moves downward and increases the opening degree of the communication passage

49

. As a result, the displacement of the compressor C is decreased. Thus, the flow rate of refrigerant in the refrigerant circuit decreases, which decreases the pressure difference &Dgr;PX.

As described above, the target value of the pressure difference &Dgr;PX is determined by the duty ratio of current supplied to the coil

64

. The control valve

43

automatically determines the position of the transmission rod

51

(the valve body

52

) according to changes of the pressure difference &Dgr;PX to maintain the target value of the pressure difference &Dgr;PX. The target value of the pressure difference &Dgr;PX is externally controlled by adjusting the duty ratio of current supplied to the coil

64

.

As shown in

FIG. 1

, the pulley

17

has an upstream pulley

17

A and a downstream pulley

17

B.

The upstream pulley

17

A includes a first outer cylinder

17

D, a first inner cylinder

17

E, and a first disk

17

F. The first outer cylinder

17

D has a power transmission portion

17

C about which the belt

18

is wound. The first disk

17

F is integrally formed with the first outer cylinder

17

D and the first inner cylinder

17

E to connect them with each other. The power transmission portion

17

C is formed on the outer circumferential portion of the first outer cylinder

17

D.

Breakable members, which are substantially columnar power transmission pins

17

G (only two are shown) in the first embodiment, are secured to the front surface of the first outer cylinder

17

D at equal angular intervals in the circumferential direction of the first outer cylinder

17

D. The power transmission pins

17

G are fit in holes formed in the front surface of the first outer cylinder

17

D. The power transmission pins

17

G project forward from the first outer cylinder

17

D and are substantially parallel to the axis of the rotary shaft

16

. The power transmission pins

17

G form shutoff mechanism for shutting-off excessive torque transmission between the engine E and the rotary shaft

16

.

The power transmission pins

17

G are made of sintered metal. The fatigue ratio &sgr;W/&sgr;B of the sintered metal is about 0.5. The sign &sgr;W represents the fatigue limit and the sign &sgr;B represents the tensile strength.

The downstream pulley

17

B includes a second inner cylinder

17

H, a second disk

17

J, and a second outer cylinder

17

K. The second disk

17

J is formed integrally with the second inner cylinder

17

H and extends radially outward from the front end of the second inner cylinder

17

H. The second outer cylinder

17

K is integrally formed with the second disk

17

J and extends rearward from the outer circumferential portion of the second disk

17

J.

Shock absorbers, which are rubber dumpers

17

L in the first embodiment, are secured to positions corresponding to the power transmission pins

17

G at the rear surface of the second outer cylinder

17

K of the downstream pulley

17

B. Each rubber dumper

17

L is accommodated in one of bores formed in the rear surface of the second outer cylinder

17

K. Each rubber dumper

17

L is cup-shaped and receives the corresponding power transmission pin

17

G.

Therefore, in the pulley

17

of the first embodiment, power transmitted from the engine E to the upstream pulley

17

A by the belt

18

is transmitted to the downstream pulley

17

B by the power transmission pins

17

G and the rubber dumpers

17

L. That is, the power transmission pins

17

G and the rubber dumpers

17

L are located in a power transmission path between the upstream pulley

17

A and the downstream pulley

17

B.

In the first embodiment, the upstream pulley

17

A, the downstream pulley

17

B, the power transmission pins

17

G, and the rubber dumpers

17

L constitute the pulley

17

. The pulley

17

has an inner space surrounded by the upstream pulley

17

A, the downstream pulley

17

B, and the like.

A substantially cylindrical hub

65

is fixed to the front end of the rotary shaft

16

. A first clutch mechanism, which is a pulley one-way clutch

66

in the first embodiment, is located between the hub

65

and the second inner cylinder

17

H of the downstream pulley

17

B. The pulley one-way clutch

66

is a first one-way clutch located in a power transmission path between the pulley

17

and the rotary shaft

16

.

The pulley one-way clutch

66

is constituted by a clutch portion

67

, and a bearing portion

68

. The clutch portion

67

and the bearing portion

68

are integrally formed with each other and arranged next to each other in the axial direction of the rotary shaft

16

.

The pulley one-way clutch

66

includes an outer ring

69

, which is secured to the inner circumferential surface of the second inner cylinder

17

H, and an inner ring

70

, which is secured to the outer circumferential surface of the hub

65

and surrounded by the outer ring

69

. The outer ring

69

and the inner ring

70

rotate relative to each other by rotating bodies, which are balls

71

in the first embodiment. The balls

71

are arranged circumferentially in line between the outer ring

69

and the inner ring

70

.

The bearing portion

68

of the pulley one-way clutch

66

, which is located between the downstream pulley

17

B and the hub

65

, is located close to the center of gravity of, or rearward of, the pulley

17

.

As shown in

FIG. 3

, recesses

72

are formed at equal angular intervals around the rotary shaft

16

in the inner circumferential portion of the outer ring

69

. A power transmission surface

73

is formed at the trailing end of each recess

72

. A roller

74

is accommodated in each recess

72

parallel to the rotary shaft

16

. Each roller

74

is movable from the position where the roller

74

is engaged with the power transmission surface

73

as shown in FIG.

3

(

a

) to the position where the roller

74

is disengaged from the power transmission surface

73

as shown in FIG.

3

(

b

).

A spring seat

75

is located at the leading end of each recess

72

, or the end of each recess

72

that is opposite to the power transmission surface

73

. A spring

76

is arranged between each spring seat

75

and the corresponding roller

74

to urge the roller

74

toward the position where the roller

74

is engaged with the power transmission surface

73

.

As shown in FIG.

3

(

a

), when the outer ring

69

rotates in the direction indicated by an arrow by power transmitted from the engine E with the pulley

17

, each roller

74

moves toward the corresponding power transmission surface

73

by the force of the spring

76

. Then, the roller

74